Hydraulic pumps and motors are well known and widely used, being one of the basic prime movers utilized by all industrial societies. These hydraulic machines are used in many different types of automotive vehicles, construction apparatus, machine tools, and manufacturing processes; and they come in all sizes, ranging from small displacement units (less than 1.2 cu.in.) to very large units (displacing more than 15 cu.in. of fluid per revolution). Among the small units are very high-speed/high-pressure devices that are only used for relatively low horsepower requirements, e.g., moving the ailerons of aircraft, positioning controls for machinery, etc.; and the large end of the range includes very low-speed/high-torque units used to operate earth movers, backhoes, etc. Many of these pump/motors have variable displacement capabilities, and the invention herein has particular relevance to such variable units which are capable of displacing more than 1.0 cu.in. per revolution. While the commercially-available pump/motors of this latter type are capable of withstanding pressures as high as 6,000 psi, structural design limitations make them generally incapable of attaining speeds over 4,000 rpm.
Further, all commercially-available variable pump/motors of which we are aware are of hydrostatic design. (Note: As used herein, the term "hydrostatic" is used to identify machines which deliver generally constant torque; and, in contrast, the term "hydrodynamic" is used to identify hydraulic machines capable of delivering generally constant horsepower.) For reasons that will be more fully explained below, the efficiency of hydrostatic pumps decreases inversely with the speed of operation, i.e., such pump/motors are only fully efficient when run at their top operational speeds. However, for most uses, it would be desirable to have pump/motors which remain efficient over a variable range of speed and torque requirements, e.g., to satisfy the continuous hydraulic power requirements of automotive transmissions.
This inefficiency problem persists in spite of the fact that the design of hydraulic pumps is an old and well-developed art. All types of pump elements relating to every aspect of pump design are well known. Prior art patents show hundreds of designs for cylinders and pistons; and, for controlling the intake and exhaust of fluid from the cylinders, they show numerous types of ball valves, spring check valves, shuttle valves, valve-plates, etc. Similarly, various designs for both fixed and variable-angle swash-plates have been disclosed for many decades in patent references; and these include split designs having a non-rotating-but-nutating portion coupled with a second portion that both rotates and nutates, the patent art showing such swash-plates being connected to drive shafts by bolts, T-bars, sliding guides, etc. Also, this same patent prior art discloses a wide variety of apparatus for controlling the adjustment of such variable swash-plates, e.g., screw threads, inclined planes, hydraulic servo mechanisms, etc.
However, in spite of all of this well-developed prior art, and in spite of the clearly indicated need for a commercially-satisfactory hydrodynamic pump (e.g., for use with vehicle transmissions), no one has created such a pump. That is, no one has been able to combine the elements of a variable swash-plate machine in a commercially-feasible structure that can satisfactorily perform under the wide range of pressures and speeds necessary for hydrodynamic operation.
While there are hundreds of patented pump designs in this old and developed art, most of these have apparently never proved successful in the marketplace. Therefore, the remaining portion of this background section will discuss only pump designs that are presently available commercially. Further, in order to facilitate appreciation of the significant improvement provided by the invention herein, the following discussion of commercially-available prior art will be generally limited to medium-sized pump units which are appropriate for a wide variety of automotive and industrial uses; and, for purposes of comparison, performance specifications will only be quoted for such pumps having generally similar displacements (approximately 4 cu.in.).
It is common for such medium-sized pump/motors to operate with maximum speeds around 3,000 rpm and maximum pressures of about 3,000-6,000 psi. While these medium hydraulic pressures can usually be contained appropriately in a relatively easy manner, containment becomes a significant problem for higher pressures, except in those few instances where other design limitations do not prevent an appropriate change in the weight or size of the hydraulic units to assure needed strength for safety or efficient sealing.
However, most industrial uses do have such design limitations. Size is particularly limited in the automotive industry where every extra pound reduces automotive efficiency and where space is at a premium. However, even in those instances where size is not a problem (e.g., large earth-moving equipment), operating pressures are limited due to cost restraints and sealing problems. Therefore, most industrial pumps are designed for maximum pressures determined by the torque requirements accompanying their desired maximum speed.
For instance, the hydraulic units used to operate vehicles are designed so that they can appropriately contain the pressures developed when the vehicle's maximum available input horsepower is applied at the pump's highest operating speed, namely, when it is being rotated 1:1 with the vehicle's engine.
Industrial pumps are similarly designed to safely contain the maximum pressures developed when the pump is delivering full horsepower at its highest specified speed. However, as the speed of such a pump is reduced and the same maximum horsepower is applied to the pump's input, the pressures within the pump increase proportionally to its reduction in speed. Since the units are designed only for the maximum pressures developed at top speed, the pressure increases developed at lower speeds must be bled off to avoid exceeding the top-speed pressure limit. This bleed-off of pressure represents lost efficiency. For example, assume that a pump is designed for a maximum pressure of 6,000 psi when providing 100 hp at a top speed of 2,400 rpm. When the pump is operated at one-quarter speed (600 rpm), its torque must be maintained, by necessity, at the limit of 6,000 psi. Therefore, at one-quarter speed, this hydrostatic unit can provide no more than 25 hp, i.e., its maximum power output at this lower speed is limited to only one-quarter of its available input power.
Mechanical drives are not so limited: In contrast, when power is supplied through a mechanical drive and a gear reduction box, the full input horsepower can be supplied at all reduced speeds; and therefore, when the speed of the mechanical drive is reduced by a 4:1 gear ratio, at one-quarter speed it provides a torque that is four times greater. Comparing this with the prior art hydrostatic device just referred to above, in order for the latter to provide the same 100 hp at one-quarter speed, the torque would have to be increased to 24,000 psi. The inability of presently-available, medium-sized hydraulic machines to operate appropriately under such pressures severely limits their efficiency. (Of course, as an alternative, it would be possible to replace the medium-sized unit with one having four times as much displacement; but such a larger hydraulic unit would cost much more, would be much larger, and would weigh more than twice as much.)
Therefore, presently-available constant-torque pump/motors are often associated with mechanical gear boxes which mechanically reduce the rotational speed of the output while maintaining the hydraulic drive at more efficient higher speeds. Although there are many power needs that could be more effectively met by full-range hydraulic machines, presently-available designs cannot produce full horsepower throughout a full range of desired rotational speeds and/or within desired limitations of size and weight.
Commercially-available hydraulic machines come in one of two basic design formats. In the most commonly used design, the pump's cylinder block rotates; and the pistons of its rotating cylinders are reciprocated as pivoting "shoes", which are attached to the end of each piston, slide over the surface of an adjustable-angle swash-plate. For controlling the flow of pressurized fluid in prior art pumps of this first basic design, the end-ports of the rotating cylinders sweep over a valve-plate, and the cylinder block must be very heavily biased against the valve-plate to minimize blowby. Therefore, the shoes and valve-plates must be regularly replaced to prevent excessive blowby and an accompanying loss in volumetric efficiency.
Further, with pumps of this first basic prior art format, as cylinder bore size is increased to meet higher power requirements, the weight and radius of the spinning cylinder blocks must necessarily be increased. This, of course, results in a proportional increase in centrifugal forces acting on the rotating cylinder blocks and pistons, placing fairly severe limits on rotational speeds. For instance, in addition to requiring more massive support structure, these increased centrifugal forces also cause the outside of the rotating piston-end shoes to lift off the surface of the swash-plate and force the extended pistons out of alignment with the cylinders, resulting in increased blowby.
Still another problem affects this first basic prior art design. Namely, the spinning cylinder blocks rotate in fluid-filled chambers; and, as speed is increased, such pumps become more inefficient due to power losses which result as their spinning cylinders move through oil. This churning of the oil increases its temperature, requiring the use of large oil reservoirs and often heat exchangers or coolers to reduce the temperature of the oil.
Because of these just-mentioned problems, it has long been recognized in the industry that it would be preferable to design machines in which the cylinders remain fixed in the housing and the pistons reciprocate against a rotating-and-nutating swash-plate. This latter design is used in the second commercially-available format. While hydraulic machines of this second design achieve higher pressures, their swash-plates are not adjustable. Thus, when used in pump/motor combinations, the speed of the motor is controlled by bleeding off volume and pressure from the pump's fluid output. In this regard, it has also been long recognized that such swash-plates should be angularly adjustable to provide speed control, but no one has been able to design such an angularly-adjustable device that is commercially satisfactory. Further, in a manner similar to the pumps of the first format, the fixed-angle swash-plates of this second format nutate in oil-filled chambers, resulting in similar churning, power losses, and temperature increases. Further, the fixed cylinders of these pumps cannot be filled rapidly, resulting in fairly severe restrictions on their practical operating speeds. Two such examples are a 4.6 cu.in. pump by Dynex-Rivett and a 4 cu.in. model by Oil Gear which, while capable of operating efficiently at 8,500 psi and 15,000 psi, respectively, have respective top speeds of 1,800 rpm and 2,200 rpm.
As indicated above, prior art patents disclose adjustable swash-plate designs using split swash-plates having two elements, namely, a non-rotating-but-nutating portion coupled with a second portion that both rotates and nutates. However, we are Unaware of any presently-available commercial pumps or motors using such prior art designs. This apparent lack of success may be related to the difficulty of providing an acceptable structure for supporting the non-rotating portion in a manner that prevents the collapse of the piston-connecting rods under the tangential forces that are created by the relative rotation between the split portions of the swash-plate. Many of these patented structures prevent rotation of the nutating-but-non-rotating portion by fixing it to a support system that includes a block sliding in a channel in the housing. Of course, with this type of restraining means, the mass of the block and its mounting must be reciprocated at high speeds against one side or the other of the channel, since the block must move back and forth for each nutation of the split swash-plate. Such structures have apparently been less than satisfactory.
A further problem that affects the efficiency and versatility of both of these commercially-available design formats relates to controlling the flow of fluid to and from the cylinders. To avoid blowby around the valve holes during the relative rotation between the pump/motor's cylinders and the valve-plate, the plates and cylinders of these pumps are heavily spring biased against each other, so the use of substantial mechanical force is required to overcome the static friction between these members in order to initiate pump operation. Therefore, such presently-available pump/motors are often incapable of developing meaningful operating horsepower at speeds of less than 500 rpm.
Efficiency is also lost because the fluid reservoirs of most presently-available pump/motors must often be maintained at pressures of at least 100 psi in order to assist the opening of fluid intake valves, to minimize cavitation problems and, sometimes, to assure the retraction of the pistons following each power stroke.
In addition to the inefficiencies and other problems referred to above, presently-available pump/motors are also relatively large in outside dimensions as well as relatively heavy in weight. Such big housings are needed by these prior art machines for supporting (a) the rotating swash-plates or cylinders, and, in those machines using adjustable swash-plates for controlling hydraulic output, for supporting (b) the means for adjusting the angle of the swash-plate. For example: A 4.1 cu.in. pump marketed by Eaton, specified to develop 100 hp at a maximum of 2,500 rpm and 3,500 psi, weighs 70 pounds and is 8.5" in diameter. However, this latter dimension does not include a 2".times.2".times.4" attachment which houses a portion of the servo mechanism that controls the angular adjustment of the swash-plate. Similarly, Volvo sells a 4 cu.in. pump (150 hp at 2,500 rpm and 6,000 psi) that weighs 132 pounds in a 7.5" diameter housing, but also has part of its servo mechanism mounted in an external attachment which extends several inches beyond the basic pump housing.
In contrast to this prior art, the novel hydraulic machine disclosed herein is a hydrodynamic device (capable of delivering constant horsepower at reduced speeds) which, for example, with a 4.4 cu.in. displacement, can provide 456 hp at 5,000 rpm and 4,000 psi. Further, these just-stated specifications can be achieved by one of our hydraulic units that weighs only 30 pounds and is contained within a housing that is only 4.875" in diameter, and that latter dimension includes the servo mechanism which controls the swash-plate.